Hydraulic valve-operating system operable to vary valve lift and timing



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t 2 W Ew C LA. B m 7 V M F. C. REGGIO HYDRAULIC VALVE-OPERATING SYSTEM OPE TO VARY VALVE LIFT AND TIMING Dec. 16, 1952 Flled Jan 2O 1943 2,621,640 ERABLE REGGIO HYDRAULIC VALVE- RATING SYSTEM O TO VARY VALVE LIFT AND TIMIN Filed Jan. 20, 1943 A Dec. 16, 1952 4 sheets-sheet 2- Dec. 16, 1952 F C REGGlO 2,621,640

HYDRAULIC VALVE-OPERATING SYSTEM OPERABLE T0 VARY VALVE LIFT AND TIMING Filed Jan. 20, 1943 4 Sheets-,Sheet I5 FCKW 2,621,640 ERABLE Dec. 16, 1952 F. c. REGGIO HYDRAULIC VALVE-OPERATING SYSTEM OP TO VARY VALV E LIFT AND TIMING 4 Sheets-Sheet 4 Filed Jan. 20, 1943 Patented Dec. 16, ,1952

UNITED STATES PATLEN T 4'O F F ICE HYDRAULIC VALVE-OPERATING SYSTEM VOPERABLE TO VARY VALVE 'LIFT .SAND

TIMING 13 `Claims. 1

This invention relates to the fluid pressureactuation'of reciprocating members and 'more particularly to mechanisms for hydraulically operating the valves of internal combustion engines and the like, and the v.plungers of engine fuel injection pumps.

'One object of the'invention is to provide alight, compact and simple valve mechanism.

Another lobject is to provide a hydraulic `mechanism permitting a considerable increase in the valve velocity,

A further object vresides 'in the Aprovision 'of va valve mechanism Autilizing lubricating `oil 4under high pressure for actuating and 'cooling the engine valves.

Still another object is to provide a hydraulic valve mechanism which permits a lconsiderable 'reduction inthe Weight and size ofthe engine.

A further object is to provide, in 'substitut-ion 'for lthe valve springs,'means for applying a fluid pressure load to the valves tending to 'close `the latter. p

A still further object is to provide-means for. automatically `altering 4the valve timing with changes of Vengine operative conditions.

Another object resides in the ,provision of --a valve vmechanism wherein the number and Weight of the reciprocating parts -is considerably reduced.

A further vobject resides in the hydraulic actuation of the plunger of a fuel injection pump.

Other-objects andadvantages -of the invention willbecorne apparent from the following description -taken in 'connection with the accompanying drawings, in which:

Fig. 1 is a fragmentary vertical section show- Iing one 'of the preferred embodiments of the linvention applied to a radial air-cooled aircraft engine.

Fig. v2 is a fragmentary section through the valve of Fig. 1.

Fig. 3 is a fragmentary front view ofthe engine Yof Fig. 1.

Figs. 4 and 5 are diagrams indicating the val-ve timing under dierent `engine operative conditions.

Fig. 6 vis -a section illustrating a secondrpreferred embodiment of the invention applied to a 'liquid-cooled engine.

Figs. 7 AandfS are sections along the :lines l-l' vand 8 5 of Fig. 6, respectively.

Fig. 9 shows a `detail of Fig.. 6 in larger scale.

Fig. `l() is a section Yillustrating a third -pre- Yferred Vembodiment 'of valve Vand fuel pump actuating mechanism according to the invention, applied to 'a liquid-cooled engine.

Fig. 1l is `a section iof ,a detail of Fig. k1-0 in larger scale.

Fig. l2 is a Adiagrammatic View `of the mechanism lof Fig. 10 when applied to a 'six cylinder in-l-ine engine.

Fig.. `13 is a 'fragmentary section through a rotary 'control valve Afor .a nine cylinder radial engine.

Fig. 14 is a'section'along line A-A of Figure l.

The'i'llustrativefexample of embodiment shown vinFi-g. 1 includes conventional parts of a'radial air cooled aircraft engine-Which will easily be recognized by those skilled in the art, namely a .portion'of cylinder head -I 3, a valve I4, and a cam '.'I'S'mounted within'the engine housing I6 'for actuating "the valves. The valve I4 is slidably 'mounted in a guide 2l above which there is pro- ;vi'ded afcylin'der ZUhaVing a'bore coaxial with the guide '21, vand cooling fins I9. The cylinder 2l), preferably 'carried by or attached to the guide 2|., `may `besecured to the cylinder head I3 by screws, not shown. A piston I8, closely iitting .the cylinder 2U and slidable therein, is attached :to the stem `of the valve VIll by vmeans of a vnut and;lock'nut. 'Ihe annular chamber 23 below the 'piston I8 is 'connected by means of port 24 -and 'pipes 25, '26 with. an .annular `conduit 23, while the cylinderchamber 3| above piston I8, closed bya cover 22, is 'connected by means `of port 22 and pipe 3'4 with the upper chamber 64 of a differential cylinder 35 having slidable therein a differential plunger v323 kactuated 'through a roller 40 from the'cam I5. The plunger has an lupper portion of smaller diameter slidable in the bore il ofthe cylinder 36, and a coaxial lower portionfof larger diameter slidable in the bore 42, thus providing anannular chamber 45 connected Vlby means of port .4S and duct 48 with the an- -nularconduit 28.

.The plunger 38 is `provided with an axial vbore 250 wherein there `is slidably Vmounted a piston .5I Aupwardly urged by a .calibrated spring 52. A -ring'53 seated in a groove .formed in the bore V5! vlimits .zthe .upward displacement of piston 5I. .The :space below the latter communicates by-way .of adrain conduit `with the engine crankcase.

An annular groove 56 formed in the plunger 38 between chambers 44 and 45 is connected by way of duct 58 with the conduit 55. A small orifice 555 may be provided in the piston 5| to establish ilow communication between chamber 44 and the drain conduit 55. An axial groove 6G formed in the lower cylindrical portion of the plunger SS communicates by way of a duct 6I formed in the latter with the chamber 44 through an orifice controlled by a ball or check valve 62 permitting oil flow from the groove 66 to the chamber 44 but preventing the return flow from said chamber back to the groove 6%. Tre latter also communicates with an annular duct 64 which receives lubricating oil under pressure from an engine pump 65 through a pipe 66, a flow-restricting calibrated orifice 68 and a pipe 10. Connected with the latter there is provided a pressure regulating plunger valve 'l2 controlling a discharge orifice 7:1 and loaded by a calibrated spring 'l5 mounted between said valve and a spring seat carried by a slidable rod 'i5 coaxial with the spring 53, whereby the load of the latter and in turn the oil pressure in the duct 64 are determined by the axial adjustment of said rod 15.

The rod 'l5 extends through the wall of a housing 'l and is attached to the movable wall of a bellows i8 contained therein and connected with a spring B3 which may be adjusted by the operator by means of a manual control 82 and a suitable linkage B3. A pipe 84 connects the engine induction manifold 85 with the space within the bellows le; and the space comprised between the housing i@ and the bellows I8 may be connected by means of a conduit 86 and a threeway cock 83 either with the pipe 9G open to the atmosphere, or with pipe ai connected with the engine exhaust manifold 92. With the cock 83 as shown in Fig. 1 the bellows 'E8 is surrounded by the exhaust pressure; and if the cock is rotated 90 clockwise the bellows 'I8 will be surrounded by atmospheric pressure.

An engine driven pump, preferably of the highpressure type having two or more stages and diagrammatically represented at 93, receives oil from a suitable reservoir, dump, or other source, for example from the discharge line 66 of the engine driven lubricating oil pump 65, and through pipe el?. supplies high pressure oil to the annular conduit 28 and to the oil accumulator and pressure regulator cylinder 95 having slidable therein a piston 9@ loaded by a spring 98. The piston controls axially spaced orifices i90 and iil through which the excess oil is discharged into a low-pressure return conduit m2 which, in the example lshown in Fig. l, is connected with the intake side of pump 93. A hand pump 103 may be provided to supply pressure oil to the accumulator S5, and thence to the annular duct 6d, through pipe E54, and prime the hydraulic system before starting the engine.

The oil discharge of the volumetric enginedriven pump 93 is substantially proportional to the engine speed. When the engine is operating within a predetermined speed range the larger orice li is covered by the piston 95 and said oil discharge minus the various leakages cf oil which are negligible as compared with the output of pump 93, is forced through the small calibrated orifice The oil velocity through this oriiice is thus Substantially proportional to the engine speed, and the corresponding drop of pressure between upstream and downstream sides of oriice it is therefore proportional to the square of the engine speed, thus maintaining the oil pressure in the annular conduit 2S at a value which is also substantially proportional to the square of the engine speed. If said speed decreases below a predetermined lower value, the piston 96 covers part of the orice lii and prevents the oil pressure from dropping below a corresponding minimum pressure limit. Conversely, if the engine speed increases beyond the maximum rated speed the piston et* uncovers the oriiice it! and prevents the oil pressure in the conduit 28 from increasing above a preselected corresponding maximum pressure limit.

in the preferred embodiment of the invention, the orifices me and le! are so spaced and the spring $28 is so designed as to obtain the following operation. When the engine is stationary, the piston Q6 is in its extreme left position and covers both orifices lili) and itil. As soon as the engine is cranked, the pump 93 discharges oil into the various lines and in the cylinder 95. Since no oil outlet is provided, the oil pressure builds up and moves the piston 35 to the right until the orifice IBD is partly opened or uncovered. As the capacity of the pump S3 is considerably larger than the various oil leakages on the discharge side of the pump even at very low engine cranking speed, as soon as the engine crankshaft is turned, the piston QS assumes such a position as to partially open the discharge oriiioe ISBS. Under these conditions, the specilic pressure of the oil on the discharge side of the pump 93 will of course be equal to the load of spring 98 (corresponding to said position of piston g5) divided by the piston area. It follows that even at the lowest cranking speed of the engine, enough oil pressure is maintained to properly operate the hydraulic valve gear, as set forth hereinafter.

As the cranking speed of the engine increases, the piston 96 moves further to the right to increase the open area of orifice Miti, until the latter is completely uncovered. In the preferred embodiment of the invention this happens just when the engine operates at idling speed. It is therefore apparent that for all values of engine speed from a few R. P. M. up to idling speed the position of the piston d varies merely by an amount equal to the diameter of the small orice lili). As a result, the load of spring 98, and the oil pressure which is proportional to the spring load, are approximately constant for all engine speeds up to idling speed. Within such a speed range the spring-loaded piston and the orifice m0 operate like a conventional pressure regulating valve, uncovering more or less of the area of orice lil@ to accommodate the variations in the delivery of the volumetric pump 93, which delivery is proportional to engine speed, while the oil pressure upstream of the orifice and the velocity of iiow of the oil through the uncovered portion of the orifice I remain approximately constant.

As already stated, at idling speed the position of piston 96 is such as to fully uncover the orifice IBG. On the other hand, the larger orifice ll is so located as to be uncovered by the piston 9G only when the engine speed exceeds the maximum designed engine speed. Thus for all values of engine speed comprised between idling and maximum speed, that is to say, throughout the normal operating range of engine speeds, the orice lili! is fully open, while the larger orice IGI is completely covered by the piston Se; and all of the oil delivered by the volumetric engine-driven pump S3 is discharged through the fully uncovcred oriiice |00. Thus for all engine speeds within said operating range the velocity of oil flow through orifice I8!) is proportional to the engine speed, and it follows of course thatthe oil pressure upstream from the orice |80, in the cylinder 95, is proportional to the square of the engine speed.

The annular chamber 23 in the valve actuating cylinderZ, connected with the conduit 28, serves to apply to-the piston I8 an upward oil pressure load tending to close the valve I4. Within the preselected engine speed range said load is proportional to the square of the engine speed and therefore proportional -to the inertia of the valve. The conventional spring or springs connected with the engine valve may thus be eliminated. In similar manner the annular chamber 45, also connected with the conduit 28, exerts on the plunger 38 a downward oil pressure load which acts against the inertia of said plunger to maintain at all times the roller 4I) thereof in contact with the cam I5.

As the engine operates, oil is displaced at each cycle to and from the small annular chambers 23 and 45, and oil pressure pulsations occur which keep the pistonV 96 of the pressure accumulator 95 in continuous motion. In an engine having a large number of valves at different phases these pressure surges are very considerably reduced. In order further to weaken the intensity of these surges, pressure cushioning means such as chambers containing a compressible fluid or chambers having yielding wall means may be provided in connection with any suitable part of the hydraulic system, for example near the valve cylinders 28 or the differential plungers 38, or the conduit 28, as it will be obvious to thosel skilled in the art. An

example of such cushioning means is illustrated in Fig. 1 in the form of a cylinder III having a chamber II4 therein connected with the conduit 28 and defined in part by a lslidable piston II2 loaded by a spring I I3. The space at the opposite side of the piston II2 is connected by means of a duct with the drain pipe Illl.

Lubricating oil leaking from the high-pressure chamber 23 between the valve stem and the guide 2I is trapped in the groove IDB and led back to the engine sump by way of the drain pipes |08 and I I0. Oil upwardly leaking from the chamber 45 is trapped in the groove 56 and returned to the sump by means of the ducts 58 and 55.

Fig. 1 shows the valve I4 in closed position, the plunger 38 in its lowermost adjustment, and the small piston 5I in intermediate adjustment between its uppermost and lowermost positions relative to the plunger 38, the load of the piston spring 52 being balanced by the oil pressure in chamber 44. As the plunger'38 is lifted by the rotating cam I5 the oil pressure in said chamber increases and displaces the piston 5I downwardly against the load of spring 52, whereupon oil is forced through pipe 34 into the chamber 3I and lifts the valve I4. The deceleration and return stroke of the latter are determined by the upward load exerted on the valve piston I8 by the high pressure oil contained in the annular chamber 23. When the valve I4 closes the plunger 38 is still in its return stroke, the oil pressure in chamber 44 drops to a low value, and the spring 52 expands and lifts the piston 5I. However during this cycle some oil has escaped through the small orice 54 and various other oil leakages have occurred, and the piston 5I will therefore attain a higher position relative to the plunger 38 than it had at the beginning of the cycle. The oil pressure in chamber 44 will accordingly be lower than in the duct 64, causing the check valve 62 to open and admit a vcertain amount of oil into chamber 44.

.It is thus apparent that the displacement of the piston 5I has the effect of rendering a portion of the stroke of plunger 38 inoperative, causing the valve I4 to open later and close earlier. If the pressure in the annular duct 64 is sufficiently low the piston 5I will attain its upper most position and rest against the ring 53 when the plunger 38 is at rest between lifts, and the valve I4 will remain open during the minimum designed duration. Conversely, if the pressure in conduit 64 is maintained at sufficiently high value so as to keep the piston 5 I- constantly in its lowermost position, the latter. becomes ineffective and the period during which the valve I4 remains open reaches its maximum designed value. By this arrangement the valve timing may be controlled while the engine is in operation as shown diagramy matically in Figs. 4 and 5, wherein b and c indicate the intake and exhaust phases respectively, and a, represents the scavenging phase during which both intake and exhaust valves are open. The former iigure, corresponding to the higher value of the oil pressure in the annular duct 64, obtained when the rod 'I5 is in its extreme left adjustment, shows a considerable scavenging phase or valve overlap, during which the cylinder may be swept by the air or mixture from the supercharger. Conversely, if the rod 'I5 is adjusted in its extreme right position; the oil pressure in duct 64 is kept at the lower designed limit, and thescavenging phase is reduced to zero or to a negligible value as shown in Fig. 5.

The bellows I8 cooperating with the spring 89 actuates the rod I5 to increase or decrease the duration of the scavenging phase or valve overlap as the pressure in the engine induction manifold becomes higher or lower than that in the exhaust pipe 32. With the three-way cock 88 rotated 90 clockwise, the bellows i8 becomes responsive to the difference between the engine manifold pressure and the surrounding atmospheric pressure, and operates the rod 'I5 to reduce the valve overlap to a negligible value when said pressure difference becomes negative. The return flow through the engine cylinder at the end of the exhaust stroke from the exhaust pipe 92 to the induction manifold 85 when the pressure in the latter is low, as when the engine is idling, may thus be avoided. The valve timing may also be adjusted by the operator while in flight by means of the control member 82.

In Fig. 2 there is indicated a valve I4 having a hollow stem or cavity II5. A rod H6 carried by the cover 22 extends therein leaving an annular interspace through which during engine operation oilflows back and forth at high velocity and thus effectively cools the valve by transferring heat by conduction and convection from the valve head to the stem portion thereof, in substantially the same way as in the sodium-cooled valves which are extensively used in aircraft engines. A helical groove III is provided in the rod IIS, and ports 24 and 32 are preferably directed tangentially with respect to the cylinder 28 as shown in Figure 14 to impart rotary motion to the oil therein and further improve the cooling of the valve II4. Oil cooling devices, such as fins I9 and 38, are provided in connection with the cylinder 28 and the oil conduits 24 and 25.

Fig. 3 diagrammatically illustrates the various hydraulic connections in the valve mechanism of a radial aircraft engine, wherein the cylindrical chamber 3I of each valve cylinder 20 is connected with the corresponding chamber 44 of a plunger housing 3G by means of a pipe 34, while the annular chambers 23 and 45 are all connected by means of pipes 25-26 and 48, respectively, with the annular conduit 28 which is provided with one or more pressure cushion chambers I and receives high pressure oil through the line 94. The mass of valve I4 with piston I8 attached thereto is only a fraction of the total mass of a conventional valve with the reciprocating mechanism connected therewith, and accordingly the load exerted on piston I8 by the oil pressure in the` annular chamber 23 need be only a small fraction of the load of the conventional valve springs. If high pressure oil is employed, the bore of cylinder 26 may be made only slightly larger than the stem of the valve I4. Furthermore, owing to the fact that only axial load is applied to the Valve and that the latter is cooled by means of lubricating oil, the valve guide 2| may be shortened. The cylinders may therefore be made very small as compared with the relatively bulky conventional valve rocker housings, and the frontal area and the weight of the engine may be considerably decreased. In addition, the cylinder head becomes more simple, and more space in the immediate vicinity of the combustion chamber becomes vavailable for cooling fins.

In the above described mechanism each valve is actuated from a corresponding cam-actuated plunger by means of `a substantially incompressible liquid column, the valve lift being determined by the cam pro'le. Fig. 6 illustrates another preferred embodiment of the invention wherein an engine-driven rotary control valve is substituted for said cam and plungers. The cylinder head |2| of a liquid-cooled, valve-in-head engine is provided with intake and exhaust valves |22 and |23. The latter valve has attached thereto a piston |24 slidable in a cylinder |25 attached to and coaxial with the guide |26 wherein there is formed a groove |28 connected with a drain duct |36. The upper end of the cylinder is closed by a threade-d cover |3| and the chambers |46 and |32 on opposite sides of the piston |24 communicate by means of ports |42 and |33, shown in Fig. 9, with the pipes |43 and |34 respectively. Similarly, the intake valve |22 carries a piston |24 slidable in a cylinder having upper and lower chambers |45 and |32 communicating with pipes |61 and |34 respectively. Said pipes |34 and |34 are connected with a liquid cooled conduit |35 communicating by means of a duct |36 with the discharge side of a high pressure engine-driven oil pump |31 diagrammatically indicated as a two-stage gear pump, and with an oil accumulator and pressure regulator |38 similar to the unit g5 and having similar functions, namely to return the excess oil to the inlet side |39 of the pump |31, to maintain the oil pressure substantially proportional to the square of the engine speed between predetermined maximum and minimum values thereof, and to cushion the pressure surges in the high-pressure system. The oil pressure in chambers |32 and |32' constantly exerts on the valves |23 and |22 a load tending to close the latter.

The pipes |61 and |43 are connected with ports |66 and |65 opening in the upper and lower parts of a cylindrical bore formed in the housing |44 of a rotary control valve having a rotor |55 therein slidably mounted on a splined portion |52 of a shaft |46 journaled in the lower valve cover |48 and carrying a pinion |53 meshing with an engine-driven gear |5I. The upper portion |53 of the shaft |46 is journaled in and extends through the upper cover |54 of the valve. In this valve arrangement, designed for a four cycle, in-line engine, the shaft |46 revolves at half crankshaft speed. The rotor |55 is provided with a groove |56 engaging a lever |51 connected with an external lever |66 whereby, by moving the latter, the axial adjustment of the rotor |55 relative to the ports |56 and |65 may be varied. A gooove formed in the central portion of the rotor provides an annular chamber |62 which by means of a conduit |63 is kept in constant communication with the discharge side of the pump |31. The upper and lower portions of the bore |45 are connected with a drain conduit |64 for leading low pressure oil back to the engine sump or other suitable reservoir.

The rotor |55 in its upper portion is provided with two grooves |68 and |10, the former communicating with the groove |62 and the latter with the low pressure drain conduit |64. As the rotor |55 revolves these two grooves |63 and |15 successively register with the port |66 thereby successively connecting the chamber |40 with the high pressure side of the pump |31 and with the drain conduit. In similar manner the lower portion of the rotor |55 is provided with two grooves |1| and |12 adapted successively to register with port |65, the former groove communicating with the groove |62 and the latter with the drain conduit |64, so that in operation the cylinder chamber |46 is also successively connected with the high and low pressure oil lines |63 and |64. The controlling edges deiinin'g said four grooves |68 and |16 to |12 are, in part at least, not parallel to the axis of the rotor but formed in such manner as to vary the duration of the intervals during which the ports |65 and |66 are connected with the high and low pressure oil lines when the axial adjustment of the rotor |55 is changed.

With the rotor |55 in the angular position shown in Figs. 6 to 8 the cylinder chambers |40 and E49 are connected by means of lines |43 and |61, ports |65 and |66, and grooves |10 and |12, respectively, with the low pressure drain conduit |54, and the valves |23 and |22 are kept closed by means of the high pressure oil contained in the annular chambers |32 and |32', respectively. As the rotor |55 revolves in the direction indicated'by the arrows, the connection between port |65 and groove |12 is interrupted and connection between said port and the groove |1i is established, and high pressure oil from the pump |31 and the accumulator |38 is admitted through the pipe |43 to the cylinder chamber |46, thus determining downward acceleration of the diierential valve piston |24. As the resultant hydraulic load applied to said piston is approximately constant during the greater part of the lift, the valve acceleration will accordingly be approximately constant.

As shown in Fig. 9, at the beginning of the valve lift only the small portion of port |42 above the d'otted line B is open, the remaining portion thereof being covered by the piston |24. However the valve velocity is low at the very beginning of the lift, and even the small eiective area of the port is sufficient for admitting the necessary volume of oil without causing any appreciable drop of pressure in chamber |40. As the lift and the velocity of the valve |23 increase, the piston |24 uncovers the port |42 and the effective area thereof attains its maximum designed value. At the same time oil is displaced from the annular chamber |32 through the port |33; and as toward the end of the downward stroke ofthe valve the piston |24 approaches and attains its lowermost position indicated in dotted line at |24 and covers the portion of said port above the dotted line C, the effective open area of the latter decreases, determining a rapid increase of oil pressure in chamber |32 and causing quick deceleration of the valve so as to reduce the impact speed of the pistonA |24 against the end wall of the cylinder |25 toa suitable value. The high pressure in chamber |48 keeps vthe valve in open position until the communication between the port |35 and the high pressure groove |1| is interrupted and the connection between said port and the low pressure groove |12 is established; whereupon the pressure in chamber |411 drops and the high pressure constantly 4existing in the annular chamber |32 causes upward displacement Aof the valve |23. Toward the end of this stroke, owing to the decreasein the effective area of port |42, they oil pressure in chamber |40 rises and causes rapid deceleration of the valve |23 so as to reduce the speed of impact of the latter against the valve seat to an appropriate value. Thereafter the valve is kept closed by the oil pressure in chamber |32 Vuntil high pressure oil is again admitted to the chamber |40 by the revolving rotor |55.

'I'he intake valve |22 is actuated' by means of the upper part of the rotor |55, controlling the port |36, in similar manner, and it is therefore regarded as unnecessary to again describe itsl operation in detail.

The oil delivery of the engine-driven pump |31 is proportional to the engine speed and exceeds the total oil displacement of the valve cylinders |25 which is also proportional to the engine speed. The excess oil, which .therefore is also proportional to the engine speed, returns to the inlet side |39 of the pump through an orice |16 which may initially be adjusted by means of a needle valve |11. It follows that the oil velocity through the orifice |16 is proportional to the engine speed, and the corresponding pressure differential between upstream and downstream sides of orifice |16 is proportional to the square of the engine speed; that is to say, the oilr in the high pressure system, including chambers |32 and conduits |35, |33 and |63, is maintained at a pressure proportional to the square of the engine speed whenever the engine is operating within a predetermined speed range. The upper limit of said range is attained when, under abnormally high engine speed, the resiliently loaded' plunger which is slidably mounted in cylinder |38 uncovers a large port |18 and thereby prevents further increase of oil pressure. The lower limit of said range is attained at a predetermined low enginer speed, for example engine idling speed, as the plunger covers the orifice |16 in part so as to reduce the effective area thereof and thereby prevents further drop of oil pressure. The actuating hydraulic load applied to the valves by means of this pressure and the positive acceleration thereof are therefore proportional to the square of the engine speed, and the valve velocity is directly proportional to the engine speed. Thus the curve representing the valve lift versus the corresponding angular position of the crankshaft is indepedent of the engine speed, exactly as if the valves `were reciprocated from a cam. En the present case, however, the valve velocity may be much higher than in the conventional valve mechanism, as the valve deceleration'and inertia are not limited by the load of the valve spring. Furthermore the mass of the valve and piston |24, the only reciprocating parts, and of the oil column movable therewith, is only a fraction of the mass of the conventional valve with the reciprocating mechanism connected therewith.

With the rotor |55 having high and low pressure grooves |38 and |15 to |12 as illustrated in Fig. 6 the valve timing and the valve overlap or scavenging phase may be varied by altering the angular position of lever |63 and thereby the axial adjustment of the rotor |55 relative to the ports |55 and |63. A speed responsive device including centrifugal flyballs |83 driven from the engine by means of gears |5| and ISI actuates a resiliently loaded sleeve |82 connected with a floating lever |33 and displaces the right end of the latter downwardly or upwardly upon an increase or decrease of engine speed, respectively. An intermediate point of lever |83 is pivoted to a rod |85 secured to a piston |85 loaded by a spring IB and slidable in a cylinder connected by means of anV oil conduit |9| with a hydraulic torque meter, not shown in the drawing, whereby the'piston |35 is displaced upwardly or downwardly upon an increase or decrease of engine torque, respectively. The left end of lever |83 is connected by means of a rod |84 with the leverr |33 so as to vary the engine valve timing in dependence upon changes of engine speed and torque, and in particular to reduce the valve overlap as the engine speed, or torque,-or both decrease.

If desired, the oil pressure in the high pressure system may be maintained constant regardless of the engine speed by closing the orice |15 by means of the needle valve |11.

The valve mechanism according to the invention may be used for actuating reciprocating valves of compressors, pumps and the like, as will be obvious to those skilled in the art, and also for actuating other reciprocating members such for example as the plungers of fuel injection pumps as illustrated in Fig. 10, wherein the cylinder head 20|? of an engine is provided with a passage 232 in which there is mounted a fuel injector 23| having a nozzle port 233 clamped on a seat 254 and adapted to intermittently discharge fuel under pressure into the cylinder. Mounted in coaxial bores formed in the injector housing 20| there are a rotatable pinion 235 provided with externaly gear teeth meshing with a fuel control rack 225 transversely slidable with respect to the axis of said pinion, and a sleeve 206 having at least one port 258 connected with a groove ZIB to which fuel is led by means of a duct 2|| from a port 2|2 adapted for connection with a fuel line 2|2. The nozzle part 253 and the sleeve 256 are tightly clamped by means of a threaded annular retaining member 2|4; A plunger 2|5, slidable inA the cylindrical cavity provided in the sleeve 255, pinion 255 and injector housing 23|, has an inclined or helical scroll 2 I6 for controlling the port 238 so as to vary the fuel discharge of the injector upon changes of angular adjustment of said plunger. A piston 2|8 of larger diameter than plunger 2|5, formed integral with or secured to the upper end of the latter, slidably lits in a cylinder machined in the upper portion of the housing 23| and closed by a threaded cover 22|, thus providing a chamber 228 and an annular chamber 230 above and below the piston 2|8, respectively.

The plunger 2|5 is shown in its uppermost position, with the piston 2|8 in contact with the cover 22|, in which the fuel pressure chamber 220 is in communication with the annular fuel groove ZIB. As the plunger 2|5 is forced downwardly, it closes the port 208 and forces fuel from the chamber 220 through the nozzle 203 into the engine cylinder. The lowermost position of plunger 2|5 is attained when the piston 2|8 reaches the lower end of the chamber 230. A longitudinal slot 222 formed in the intermediate portion of the plunger 2|5 is engaged by the inner end of a pin 223 forced into a radial bore formed in pinion 205 through a suitable opening provided in the housing 20|, not shown in the drawing, whereby the angular adjustment of the plunger and in turn the fuel discharge of the injector are determined by the axial adjustment of the control rack 224, as it is well known in the art. The injector 20| is secured to the engine by means of fastening means such as screws 225.

The chambers 228 and 230 are connected by means of ports 233 and 231 which may be similar to the ports |42 and |33 of Fig. 9, with the pipes 23| and 232 respectively. The latter pipe is connected with the high pressure oil conduit |35', similar to the conduit |35 of Fig. 6, whereby an upwardly directed pressure load is constantly exerted on the piston 2|8. The former pipe 23| is connected with a port 234 opening in the bore 235 of a rotary control valve 236 including a rotor 238 slidably mounted on a splined shaft 240 journaled in the lower cover 24| and driven from the upper end |53 of the shaft |46 (shown in Fig. 6) of the control valve |44 by means of a coupling 242. The rotor 238 has a groove 243 constantly communicating with the high pressure oil conduit |63', similar to conduit |63 of Fig. 6. A cover 244 closes the bore 235 at the upper end thereof, and a low pressure drain conduit 245 is connected with the upper and lower ends of bore 235. The rotor 238 has two grooves 246 and 248 communicating with the high and low pressure lines |63 and 245 respectively, and as the rotor 238 revolves said two grooves successively register with the port 234 and cause the plunger 2|5 to reciprocate in substantially the same manner as has been described in detail in connection with the valves |23 and |22. As the controlling edges of the grooves 246 and 248 are not parallel to the axis of the rotor 238, the injection timing may be varied by altering the axial adjustment of said rotor relative to the port 234, To that end the rotor 238 is provided with a groove engaged by a lever 253 connected with an external lever 254 actuated by way of the rod 255.

The valves |22 and |23 in Fig. 10 are the same as in Fig. 6, while the valve actuating cylinder 260' or 260, the latter of which is shown in section in Fig. l1, has an upper port 26| including a main circular portion and a comparatively very small triangular portion so designed that when the upper edge of the valve piston |24 approaches its uppermost position indicated by the upper dotted line D, the eiective open area of the orice 25| becomes sufliciently small to cause quick deceleration of the valve before the latter reaches its seat, thus rendering possible the adoption of valve velocities many times as great as those employed in conventional high-speed engines. However with such form of port the initial part of the valve lift may be slowed down by the throttling effect oi the small initial effective area of the port. To obtain a high value of deceleration prior to the seating of the valve as well as unrestrained initial valve lift acceleration, a second communication is provided between the pipe |43 and the upper part of the cylinder chamber 40 by means of conduits 262, 263 and 265, including a check valve 264 permitting flow of oil from the pipe |43 into the upper chamber but preventing flow of oil in the opposite direction. A similar arrangement including a check valve 265 is pro vided at the lower end of the annular chamber in connection with the lower port 268, whereby a high value of deceleration may be obtained at the end of the valve lift as well as unrestrained acceleration at the beginning of the valve closing; stroke.

The sections of the rotary control valve |44 in Figs. 7 and 8 indicate six equally spaced radial. ports in a plane perpendicular to the axis of the rotor |55. The latter rotates in the direction of the arrows at half crankshaft speed. This arrangement is intended for a six cylinder fourcycle engine. Fig. 12- diagrammatically indicates the hydraulic connections, by means of pipes |61 and |43, between the rotary control valve |44 (the upper fragment of which is shown in Figure 10) and the intake and exhaust valves 260" and 260, respectively, as well as the connections, by means of pipes 23|, between the rotary control valve 236 and the fuel injectors 20| of a six-cylinder engine 255 wherein the ignition order is |-4-2-6-5--5. It will be noted that each cylinder is provided with two exhaust valves 260 connected with the rotary valve |44 by means of a common pipe |43, and with two intake valves 260 connected with said rotary valve by way of a single pipe |61.

It will be obvious to those skilled in the art that a valve mechanism as shown in Fig. 1, wherein the Valves are operated by means of cam-actuated plungers, may be used in connection with an in-line engine as well as with a radial engine, and conversely the valve mechanisms illustrated in Figs. 6 and 9 to 1l, employing a rotary control valve, may be applied to a radial engine as well as to an in-line engine. Fig. 13 indicates a transversal section through the upper portion of a rotary control valve 210 for a nine-'cylinder air-cooled radial aircraft engine,`having nine equally spaced radial ports 21|, 212, 213, etc., connected with the intake valves of the engine cylinders. Said ports are controlled by a rotor 215 driven by a splined shaft 216 at one-eighth crankshaft speed in anticlockwise direction. The rotor 215 is provided with four equally spaced grooves 218 connected with the high-pressure oil system, alternating with four grooves 285 communicating with the oil drain line. The rotary valve is otherwise substantially similar to valve |44, and it is therefore regarded as unnecessary again to describe it in detail.

The vari-ous cooling arrangements and oil pressure regulating and pressure surge cushioning devices disclosed in connection with any of theI above embodiments may obviously be employed in connection with the others. Furthermore any suitable iluid may be used as hydraulic medium in substitution for the lubricating oil. And it is to be expressly understood that the invention is not limited to the specific embodiments shown.

gessleinj 13 but maybe used in various other ways, and various'modifications may be made to suitdifierent requirementsand that changes, substitutions, additions and' omissions may be made in the construction, arrangement and manner of adjustment. and loperation within the limits or scope of the invention-as defined in the following claims. Certain features ,disclosed herein are claimed in application- Serial No.-3l9,577 filed November 8, 1952 by the same applicant.

Where claims vare directed to` less than all the elements of the complete system disclosed, they are intended to cover possible uses of the recited elementsin installations which may lack the nonrecited elements.

What I claim is:

1. A reciprocable engine distribution member which moves at every engine-cycle, means includlng an engine-driven fluid pump and a calibrated orifice on the discharge side of the pump for exerting on said member a fluid pressure load in one direction proportional to the square of the engine speed between preselected upper and lower limits of said speed, and means for applying to said member an intermittent actuating load directed against said fluid pressure load.

2. A reciprocable valve member having piston means, first fluid containing means for continuously exerting thereon a fluid pressure load in one direction, means including cooperating calibrated orifice means and a volumetric fluid pump to vary said fluid pressure load substantially in proportion to the square of the number of reciprocations per unit time of said member whereby said load varies substantially in proportion to the inertia of' said member, second fluid containing means for intermittently applying a larger` fluid pressure load in the opposite direction to said member for reciprocating the latter, a low pressure fluid system, and valve means for admitting fluid from said system to said second containing means to compensate for fluid leakage therefrom.

3. A reciproca'ble engine distribution member which moves at every engine cycle; means to actuate said member in one direction; first wall means movable with said member; fluid conduit means to exert variable fluid pressure on said wall means in the opposite direction; a calibrated orifice; an engine driven volumetric pump discharging fluid through said orifice to maintain in said fluid conduit means a pressure substantially proportional to the square of engine speed; second movable wall means subject to the fluid pressure in said fluid conduit means; and means for exerting on the second wall means a resilient load directed against said fluid pressure to decrease the intensity of pressure variations in said fluid conduit means upon motion of said member.

4. An engine valve having first wall means movable therewith to actuate the valve; a fluid pressure chamber on one side of the wall means to exert on the latter a pressure load tending to close said valve; second movable wall means capable of a volumetric displacement comparable to that of said first Wall means and subjected to the pressure of the fluid in said chamber; means for applying to said second wall means a resilient load directed against said fluid pressure; and means for substantially increasing or decreasing the fluid pressure in said chamber upon increase or decrease of engine speed, respectively.

5. A reciprocable engine Valve, a cylinder, piston means slidable in said cylinder for actuating said valve, high pressure fluid containing means 'f4 in constant communication with said cylinder for exerting onsaid piston means a` fluid pressure load tending tov closesald valve, and means including an engine-driven fluid pump and a calibrated orifice respectively upstream 4and downstream from said fluid containing means for increasing or decreasing the pressure in said fluid containing means with an increase or decrease of engine speed respectively.

6. Engine valve mechanism. including means for varying the valve timing, and engine induction and exhaust pressure responsive means for actuatingisaidfirstn mentioned means.

7.`Engine valve actuating mechanism having means for controlling the valve overlap, and means for actuating said first mentioned means to decrease said overlap as the difference between the engine manifold pressure and the exhaust pressuredecreases.

S. R'eciprocable tappet means actuated froma rotative cam, means including a source of pressure fluid whose flow increases with the speed of said cam for continuously exerting a high pressure fluid load on said tappet means to force the latter against said cam and means including fixed-area orifice means in the path of said fluid flow for increasing said pressure fluid load with increase of speed of said cam.

9. A hydraulic valve mechanism including a reciprocable valve, a rotative cam, a plunger actuated from said cam for hydraulically lifting said valve, means including a source of pressure fluid whose flow increases with the speed of said cam for continuously exerting a pressure fluid load on said valve and plunger tending to close the former and to force the latter against said cam, and calibrated orifice means in the path of the flow of said fluid to increase the fluid pressure and thereby the load applied to said valve and plunger substantially in proportion to the square of the cam speed.

10. An engine valve, pressure responsive means exerting a continuous load on said valve in the direction to cl-ose said valve, means for cyclically lifting said valve, a volumetric fluid pump driven from said engine, a fluid flow restricting orifice having fixed flow area on the discharge side of said pump to determine therein a pressure proportional to the square of the engine speed Within preselected speed limits, said pressure TBSDOD- sive means being subject to the fluid pressure on the discharge side of said pump, whereby the continuous load exerted on said valve varies proportionally to the square of the engine speed and thereby in direct proportion to the inertia of said valve.

11. A slidable valve having a closed position, a mechanism including rotating means to open said valve periodically, fluid pressure responsive means for constantly applying a load to said valve tending to close the latter, and means including a fluid pump delivering a flow of fluid substantially proportional to the speed of said rotating means and fluid flow restricting orifice means having calibrated flow area in the path of said fluid flow to vary the fluid pressure operating on said pressure responsive means in proportion to the square of the speed of said rotating means, whereby the valve closing load varies in proportion to the valve inertia.

12. Engine v-alve mechanism including valve timing control means, and engine exhaust pressure responsive means actuating said control means.

13. Engine valve mechanism including valve timing control means, and engine exhaust pressure responsive means and manual control means actuating said timingcontrol means.

FERDINANDO CARLO REGGIO.

REFERENCES CITED The following references are of record in the le of this patent:

UNITED STATES PATENTS Number Name Date 885,459 Engler Apr. 21, 1908 1,316,977 Ricardo Sept. 23, 1919 1,474,794 Shephard Nov. 20, 1923 1,529,201- Meredith Mar. 10, 1925 1,696,984 'Irbojevich Jan. 1, 1929 1,787,120 Noble Dec. 30, 1930 1,798,938 Hallett Mar. 31, 1931 1,841,337 Roessler Jan. 12, 1932 Number Number 16 Name Date Jardine Apr. 3, 1934 Leveque Mar. 12, 1935 Lundh Aug. 20, 1935 Alden Jan. 14, 1936 Alden Feb 9, 1937 Wurtele Mar. 2, 1937 Duncan June 21, 1938 Pierce Feb. 7, 1939 Duncan May 23, 1939 Duncan Feb. 27, 1940 Alden Dec. 3, 1940 Pierce et a1 Aug. 17, 1948 Steiner Sept. 14, 1943 FOREIGN PATENTS Country Date Great Britain of 1903 Great Britain Apr. 5, 1917 France June 26, 1924 

